Copyright © 2010 KSAE 1229−9138/2010/051−01 International Journal of Automotive Technology, Vol 11, No 2, pp 139−146 (2010) DOI 10.1007/s12239−010−0019−z COMPARISON OF TWO INJECTION SYSTEMS IN AN HSDI DIESEL ENGINE USING SPLIT INJECTION AND DIFFERENT INJECTOR NOZZLES J BENAJES , S MOLINA , R NOVELLA , R AMORIM , H BEN HADJ HAMOUDA and J P HARDY 1) 1)* 1) 2) 1) 2) Universidad Politécnica de Valencia, CMT-Motores Térmicos, Valencia 46022, Spain Renault S.A.S., Rueil Malmaison 92508, France 1) 2) (Received 28 January 2008; Revised July 2009) ABSTRACT−The demand for reduced pollutant emissions has motivated various technological advances in passenger car diesel engines This paper presents a study comparing two fuel injection systems and analyzing their combustion noise and pollutant emissions The abilities of different injection strategies to meet strict regulations were evaluated The difficult task of maintaining a constant specific fuel consumption while trying to reduce pollutant emissions was the aim of this study The engine being tested was a 0.287-liter single-cylinder engine equipped with a common-rail injection system A solenoid and a piezoelectric injector were tested in the engine The engine was operated under low load conditions using two injection events, high EGR rates, no swirl, three injection pressures and eight different dwell times Four injector nozzles with approximately the same fuel injection rate were tested using the solenoid injection system (10 and 12 orifice configuration) and piezoelectric system (6 and 12 orifice design) The injection system had a significant influence on pollutant emissions and combustion noise The piezoelectric injector presented the best characteristics for future studies since it allows for shorter injection durations and greater precision, which means smaller fuel mass deliveries with faster responses KEY WORDS : Diesel engines, Injection system, Nozzle, Pollutant emissions NOMENCLATURE and particulate matter, from petrol and diesel engines Air quality, especially in large urban areas, has been impacted by engine exhaust gases and particulate matter To reduce this impact, a stricter emission regulation, Euro 5, began in 2009 Since 1996, when Euro started to tighten emission limits, new technologies have been developed to meet the challenge of complying with low emission limits In the future, regulations are predicted to become stricter for engine manufacturers In the case of Euro 5, there is already a noticeable reduction in NO and particulate matter compared to Euro It is expected that diesel particulate filters will be mandatory for all diesel cars by 2011 Euro 6, which will probably start being enforced in 2014, will significantly lower NO emission limits from the current 0.180 g/km (Euro 5) to 0.080 g/km This will force car manufacturers to invest additional resources into research, and thus increase the final price of vehicles (EurActive, 2004; The European Commission, 2006a, 2006b) Part of the Diesel engine’s great advances in performance and control of pollutant emissions over the last decade can be attributed to improvements in the injection systems The introduction of the high pressure common-rail injection system has allowed for better control of the combustion process through flexible, more accurate control of the injection parameters This system allows the number of injection pulses, the time interval between them, the injection duration and the injection pressure (IP) to be precisely ATDC : after top dead center BGT : burned gas temperature BSFC : brake specific fuel consumption BTDC : before top dead center CAD : crankshaft angle degrees CO : carbon monoxide CO : carbon dioxide EGR : exhaust gas recirculation EOI : end of injection FSN : filter smoke number HC : hydrocarbons HRL : heat release law IMEP : indicated mean effective pressure MD : mass distribution NO : nitrogen oxides RoHR : rate of heat release SFC : specific fuel consumption SOI : start of injection X X X INTRODUCTION Many health and environmental problems have been attributed to pollutant emissions, mainly NO (Nitrogen Oxides) X * Corresponding author e-mail: samolina@mot.upv.es 139 140 J BENAJES controlled In newer injection systems, it is also possible to control the injection rate shape (Robert Bosch GmbH, 2004) Despite the fact that this system increases control of the fuel injection process, it makes finding the optimum operating conditions more difficult (Desantes , 2007) Due to the need to reduce emissions, extensive research has focused on in-cylinder control of pollutant formation It is well-known that reducing NOX, smoke and HC (Hydrocarbons) emissions at the same time is a very difficult task Some strategies focus on split injections as a way to control emissions Nehmer and Reitz studied the effects of rateshape and split injection on diesel engine performance and emissions They observed that the amount of fuel in the first injection affected the engine-out emissions and incylinder pressure rise rate, which are directly related to combustion noise Higher NOX emissions and lower smoke production were seen when more fuel was injected in the pilot injection (Nehmer and Reitz, 1994) Tow investigated the effects of multiple injections on combustion in heavy-duty Diesel engine operation at medium and low load conditions Multiple injection strategies reduced NOX emissions, and the dwell time between injection events was shown to heavily influence combustion process control (Tow , 1994) Pierpont, Montgomery and Reitz tested multiple injection strategies involving EGR (Exhaust Gas Recirculation) in order to reduce NOX emissions without significant penalties on smoke and BSFC (Brake Specific Fuel Consumption) They observed that multiple injections could effectively reduce particulate matter, NOX and combustion noise They pointed out that the undesirable EGR collateral effects of increased particulate emissions might be compensated for by the use of multiple injections (Pierpont , 1995) Montgomery compared the behaviors of different nozzles in relation to the flow exit area and number of orifices and highlighted their influence on combustion According to their study, nozzles with shorter spray penetrations produced more particulate matter and lower NOX (Montgomery , 1996) Benajes also investigated the influence of nozzle orifice number and the use of swirl in a retarded split injection on gaseous emissions and combustion noise They remarked that the low temperature combustion obtained with a late injection is able to provide ultra-low NOX emissions and reasonable combustion noise at medium load conditions Their results also showed that a high orifice number is prone to causing very high smoke emissions due to an undesirable interaction among the fuel spray jets, which could be intensified by swirl (Benajes , 2006) The use of multiple injections in a small diesel engine was also discussed by Hotta They investigated how an early pilot, close pilot and post-injection could affect the combustion process and pollutant emissions In their work, they observed that large early-pilot injections could increase HC emissions due to a cylinder wall impingement It was also found that the use of post-injection helped reduce et al et al et al et al et al et al et al et al et al et al smoke emissions However, this phenomenon had previously been noticed (Hotta , 2005) Desantes et al studied the usage of post-injections They concluded that post-injections were capable of reducing smoke emissions considerably with no penalty on NOX emissions Their study was focused on post-injection and the phenomenon of soot oxidation The results revealed that the post-injection reduced soot However, it was observed that post-injections did not interact with the main injection Consequently, soot would not be reduced by enhanced soot oxidation caused by the post-injection Furthermore, engine-out soot would be the sum of the soot resulting from the combustion of the main and pilot injections separately Thus, the final level of soot decreased because the main pulse produced less soot and the post-injection did not produce significant additional soot (Desantes , 2007) Finally, Benajes carried out an investigation using a small single-cylinder engine based on a statistical procedure called “Consecutive Screenings”, which showed significant improvements in pollutant emissions by substantially increasing the EGR rate, retarding the injection event and using variable dwell time (Benajes , 2007) The objective of this work was to compare the pollutant emissions using a piezoelectric injection system and a solenoid injection system in a light duty engine with a low compression ratio of 14:1 at constant SFC (Specific Fuel Consumption), using split injections and running in a low load engine mode The influence of dwell times and split injection mass distribution were also studied, in order to evaluate the possibilities of each injector in each case Usually, in other studies, the influences of pilot- and postinjections on pollutant emissions are investigated, taking into consideration that the use of split injections can affect combustion efficiency and IMEP (Indicated Mean Effective Pressure) In this work, the analysis the analysis was performed with the amount of injected fuel and IMEP held constant et al et al et al et al EXPERIMENTAL FACILITY AND EQUIPMENT The engine used in this work was a single-cylinder research engine with a displacement volume of 0.287 liters, four valves and low compression ratio, equipped with a common rail injection system This engine corresponds to a 1.2-liter, 4cylinder engine The engine was installed in a fully instrumented test cell with all of the required facilities for the operations and control of the engine The required boost pressure (BP) was provided by a screw compressor, and the intake air was heated to 40oC Exhaust gas recirculation (EGR) was kept constant at 120oC NOX , CO (Carbon monoxide), HC, CO2 (Carbon Dioxide) and O2 (Oxygen) measurements were performed with a HORIBA 7100D gas analyzer Smoke emissions were measured with an AVL 415 variable sampling smoke meter, which provided results directly in FSN COMPARISON OF TWO INJECTION SYSTEMS IN AN HSDI DIESEL ENGINE USING SPLIT INJECTION Figure Engine experimental laboratory set-up (filter smoke number) In-cylinder pressure was measured with a piezoelectric transducer, and additional information, such as IMEP and combustion noise, could be evaluated during the tests using this data The experimental set-up is presented in Figure Combustion diagnosis software was used to calculate the heat release (HRL), rate of heat release (RoHR), burned gas temperature (BGT) and other valuable information Data recorded from 50 consecutive engine cycles with a resolution of 0.2 crank angle degrees (CAD) was used for this calculation The model is based on the solution of the energy conservation equation in the cylinder, with the assumption of uniform pressure and temperature over the instantaneous volume This single-zone model enables the calculation of the instantaneous average temperature in the burned gas, as well as the heat released during the combustion (Lapuerta , 1999) and (Desantes , 2004) et al et al METHODOLOGY The tests were carried out in two stages for each fuel injection system used in this work The first stage was to define the engine’s operational conditions The second stage consisted of the tests comparing dwell time and mass distribution 3.1 Engine Operational Condition for the Preliminary Tests In the preliminary tests, a 12-orifice nozzle was used for the solenoid system and a 6-orifice nozzle was used for the piezoelectric system Both nozzles had a conical orifice shape and similar hydraulic mass flow The engine was operated at 1500 rpm and 4.0 bar IMEP, and the dwell time was set to be 1.0 ms In this study, the dwell time is considered to be the time interval between the end of the first injection event and the beginning of second injection event Ranges for injection pressure and EGR rate were defined with the aim of performing a parametric study, and final values are shown in Table The mass distribution (MD) of 141 Table Engine operating conditions of the preliminary tests and pollutant emission targets Engine operating conditions Solenoid Piezoelectric Engine speed 1500 rpm 1500 rpm IMEP 4.0 bar 4.0 bar SOI f(SFC, IMEP) f(SFC, IMEP) Dwell time 1.0 0.6 Injection pressure 600~1200 bar 900~1300 bar EGR rate 40%~50% 40% Nozzle 12 holes holes Mass fuel 8.0 mg/cc 8.0 mg/cc MD 50%-50% 20%-80% to 80%-20% Smoke, noise and pollutant emission targets Smoke < 2.00 FSN Noise < 80.0 dB NOX < 0.25 g/kWh CO < 6.50 g/kWh HC < 1.50 g/kWh the split injection is presented in this format: 50/50 MD This nomenclature indicates that 50% of the fuel mass is injected in the first injection event and the remaining 50% is injected in the second event The engine mode and pollutant emission targets are based on the EURO cycle for a passenger diesel vehicle with an aftertreatment particulate filter (Table 1) The presence of the particulate filter in the exhaust allows for a high level of engine-out smoke emissions, as can be seen in the Table 3.2 Engine Test The engine test stage was characterized by sweeping the dwell time for various mass distributions The engine tests for both fuel injection systems were performed with two different nozzles, varying only the number of orifices Although the nozzles differ in the number of orifices (10 and 12 orifices), the theoretical hydraulic flow is very similar The ISFC was fixed at 250 g/kWh and an IMEP of 4.0 bar was targeted It is important to point out that the SOI (Start of Injection) was varied in order to keep the IMEP at 4.0 bar Engine tests using the solenoid and the piezoelectric injection systems were carried out with an injection pressure of 900 bar and a 45% EGR rate, based on the results obtained in the definition phase The boost pressure was set at 1.2 bar Using the solenoid system, the mass distribution ratio ranged from 30/70 to 50/50, and the dwell time was swept from 0.6 to 1.6 ms using both 12-orifice and 10-orifice nozzles Dwell times shorter than 0.4 ms would be very unstable, and thus were not tested Using the piezoelectric 142 J BENAJES injection system, the mass distribution ratio of the split injection was swept from 20/80 to 50/50, and the dwell times ranged from 0.2 to 1.4 ms The piezoelectric injection system responded faster and more accurately, allowing for shorter injection durations and dwell times The same tests were repeated with a constant SOI, instead of a constant IMEP, in order to separately evaluate the effects of different engine parameters on engine behavior RESULTS AND ANALYSIS 4.1 Solenoid Injection System Preliminary Test Results In the preliminary test phase, some engine tests were carried out to select the most suitable EGR rate and injection pressure for the next part of this study The results and the main observations of this part of the study are presented in Figure 2(a) and (b) Graph (a) shows the pollutant emissions for each tested case Graph (b) shows HRL, RoHR and BGT curves for 900 bar IP EGR swept and 45% EGR rate injection pressure swept As seen in Figure 2(a), for 600 bar IP, the levels of CO and HC emissions were much higher than the targeted values Unexpectedly, the NOX emissions did not seem to be related to the injection pressure It is confirmed by the BGT graph in Figure 2(b) that the burned gas temperature Figure (a) Noise, soot and pollutant emissions from the preliminary tests using the solenoid injection system; (b) HRL, RoHR and BGT vs Crankshaft angle for solenoid preliminary tests et al peak did not change significantly with injection pressure, while it decreased considerably with increasing EGR rate Smoke emissions for the lowest IP (blue circle) were under the target value but very close to the limit Smoke was also higher than observed for the other injection pressures due to worse mixing conditions This was not a good result since it did not leave margin to work on a possible tradeoff For an injection pressure of 1200 bar, the obtained CO values were slightly higher Combustion noise (in the green circle) was also considered unsuitable, so the small reduction in smoke emissions does not justify its use in the next phase Using an injection pressure of 900 bar resulted in smoke emissions below the target value Moreover, this configuration presented lower CO and HC emissions (dotted red lines) than the other IP values Thus, 900 bar seemed to be the most reasonable injection pressure for further development Of the tested EGR rates (40%~50%), 40% had the highest NOX emissions because the higher O2 concentration led to higher in-cylinder temperatures Thus, NOX formation (black arrow) was not inhibited enough to stay under the target value On the other hand, increasing the EGR rate to 50% (grey arrows) significantly reduced the O2 concentration, enough to efficiently reduce the in-cylinder temperatures and combustion noise However, this EGR rate had unreasonably low combustion, which resulted in high CO emissions Furthermore, it was concluded that the next engine tests using the solenoid injection should be carried out with the IP and EGR rate set to 900 bar and 45%, respectively In Figure 2(b), the BGT graph shows that the maximum burned gas temperature did not vary as a function of injection pressure, but decreased when the EGR rate increased In order to maintain an IMEP of 4.0 bar, the SOI had to be advanced when either injection pressure or EGR rate increased The effect was that the RoHR of the first combustion was considerably reduced (see red arrows) At an IP of 1200 bar, it seemed that the combustion of the pilot and main injections started almost simultaneously, increasing the RoHR slope and justifying the high values of combustion noise observed At a 50% EGR, the RoHR did not rise because increasing the EGR rate also caused a significant reduction in the mixing rate and, consequently, in combustion velocity (see brown arrows) 4.2 Engine Preliminary Test Results Using Piezoelectric Injection System The preliminary tests using the piezoelectric injection system intended to define an appropriate injection pressure and mass distribution range for the next engine tests This injection system allows for injection durations as short as 140 µ s, which permits injecting very small amounts of fuel, such as 20% of the total injected mass (1.6 mg/cc) Thus, the split injection was swept from 20/80 to 80/20 of the mass distribution with a fixed dwell time of 0.6 ms (5.4 CAD) There are two different ranges to be considered and COMPARISON OF TWO INJECTION SYSTEMS IN AN HSDI DIESEL ENGINE USING SPLIT INJECTION 143 Figure Noise, soot and pollutant emissions from the preliminary tests using the piezoelectric injection system observed separately: pilot injection (from 20/80 to 50/50) and post injection (from 60/40 to 80/20) Analyzing the results at 1300 bar of IP in Figure 3, the pilot injection range presented higher levels of HC and CO emissions than other injection pressures However, the emissions tended to decrease when the fuel mass of the first injection event (green arrow) was increased However, the combustion noise increased in the same range because more fuel mass was burned in the premixed combustion In the post-injection range, it is possible to observe very high combustion noise, up to unacceptable levels At 1100 bar of IP, the smoke emissions stayed at very low levels, even though the HC and CO emissions were high when using a pilot injection The smoke emissions were slightly decreased, whereas CO and combustion noise did not increase much The use of a 900 bar IP post-injection significantly reduced smoke formation, as previously known, although HC and CO emissions increased (blue circles) (Desantes , 2007; Han , 1996) Moreover, the use of a pilot injection at 900 bar of IP kept the CO and HC emissions at a lower level than the other injection pressures Although smoke emissions were higher than the other pressures, they were still very far below the proposed target Based on this analysis, the engine tests using the piezoelectric injector were carried out with a 900 bar IP and a pilot injection The EGR rate was increased to 45% in order to reduce NOX formation since some of the test points did not meet their targets et al et al 4.3 Engine Tests Analysis The engine tests with the solenoid injector were performed with a 10-orifice and 12-orifice nozzle The 12-orifice nozzle tests had a range of mass distribution from 30/70 to 50/50 MD The same tests were repeated using the 10-orifice nozzle, except for 30/70 MD because a high level of combustion instability was found when using the 12-orifice nozzle under those conditions The cause was the short Figure Comparison of (a) 40/60 and (b) 50/50 mass distributions injection duration required to inject only 30% of the injected mass, which caused the injector needle to pulse as fast as possible For both nozzles, the dwell time ranged from 0.6 to 1.6 ms Engine tests with the piezoelectric injector were carried out with a 6-orifice and 12-orifice nozzle The mass distribution range was swept from 20/80 to 50/50, and the dwell time ranged from 0.2 ms to 1.4 ms Figure 4(a) and (b) present a comparison among all the nozzles used in this study, independently of the injection system, with mass distributions of (a) 40/60 and (b) 50/50 A 45% EGR rate was used and the IMEP was isolated at 4.0 bar The dashed lines are the proposed targets for each pollutant It is important to make clear that 0.2 and 0.4 ms of dwell time were not tested with the solenoid system and 1.6 ms was not tested with the piezoelectric system The 20/80 and 30/70 mass distribution graphs are not shown here due to space limitations Increasing the dwell time contributed to smoke formation (blue circles) when using the 6-orifice or 10-orifice nozzles, as seen in Figure 4(a) and (b) The different levels of smoke emissions depend on the number of orifices of each nozzle; the 12-orifice nozzles showed lower smoke emissions than the nozzles with fewer orifices When using 12-orifice nozzles, more advanced SOI’s were necessary to reach a 4.0 bar IMEP, and they presented higher combustion noise 144 J BENAJES NOX and smoke levels did not represent a problem since they stayed below the target in the majority of the tested points However, combustion noise, HC and CO emissions did not fulfill the required limits with both nozzles and all mass distributions In general, the 10-orifice nozzles produced more smoke than the 12-orifice nozzles However, the 12-orifice nozzle presented more combustion noise because it required a slightly advanced injection timing to maintain a 4.0 bar IMEP As seen in Figure 4(b), the 6-orifice nozzle with reduced dwell times, such as 0.2 or 0.4 ms, presented unsuitable levels of combustion noise independently of the mass distribution (black arrow) NO emissions stayed below the limit The combustion noise target was achieved for both mass distributions between 0.6 and 1.0 ms of dwell time The 20/80 and 30/70 mass distributions (not shown in the picture) stayed close to the target However, HC and CO emissions are still very high in all the cases Finally, for the 50/50 mass distribution, the smoke emissions were very close to the limit of 2.0 FSN Figure 5(a) and (b) present two cases in which the configurations using the piezoelectric injector and solenoid injector were equal In Figure 5(a), both injection systems were tested using 12-orifice nozzles, 45% EGR, 4.0 bar X et al IMEP, 1.0 ms dwell time and −11.0 CAD ATDC SOI () (b) had the same configuration, except the dwell time and SOI were changed to 1.2 ms and −11.5 CAD ATDC SOI, respectively The injector opening timings of the second injection event are represented in the graphs by the vertical lines Although the engine test configurations were exactly the same in each graph, some differences are noticeable when comparing the HRL, RoHR and BGT curves for both injection systems The shorter time that the piezoelectric injector required for opening and closing caused an advance of the second injection relative to the solenoid injector Consequently, the combustion process was advanced Examining Figure 5, it can be seen that cool flame reactions started before the second injection when using the solenoid injector However, those cool flame reactions were not seen when the piezoelectric injector was used The second injections using the piezoelectric injector avoided the cool flame reactions In this case, the cool flame reactions were responsible for the temperature increase before the combustion process Furthermore, it can be seen that the maximum BGTs are very similar for both injectors, but the average temperatures during the combustion were slightly higher using the solenoid injector, which could have increased NO formation in the beginning and prolonged smoke oxidation at the end of the combustion process This effect is stronger for 1.0 ms of dwell time than for 1.2 ms Figure shows the graphs for HRL, RoHR and BGT for different dwell times corresponding to the tests with 4.0 bar of IMEP, using the 12-orifice nozzle with a mass distribution of 50/50 It can be observed that the combustion occurs more smoothly when the dwell time is increased due to retardation of the center of the combustion With a dwell time of 0.2 ms, the combustion is very similar to that of a single injection When the dwell time is increased to 1.4 ms (12.6 CAD at 1500 rpm) the combustion of both injections seems to be slower, and the premixed combustion less abrupt Dwell times longer than 1.0 ms reduced the slope of the X Figure Comparison of HRL, RoHR and BGT vs crankshaft angle curves for 40/60 MD using solenoid and piezoelectric injectors with 1.0 ms and 1.2 dwell times The engine configurations for each dwell time case were exactly the same, including injection timing Figure HRL, RoHR and BGT vs crankshaft angle for 50/50 mass distributions COMPARISON OF TWO INJECTION SYSTEMS IN AN HSDI DIESEL ENGINE USING SPLIT INJECTION 145 Figure HRL, RoHR and BGT vs crank angle for Iso-SOI tests with 1.0 ms of dwell time Figure Pollutant emissions from the Iso-SOI tests RoHR, causing less combustion noise The maximum BGT was reduced by increasing the injection dwell time But the time period that the BGT remained at NOX formation temperatures increased with increasing injection dwell time (Akihama , 2001) Thus, there was not a significant change in NOX emissions In order to study the isolated effects of mass distribution and dwell time using the 12-orifice nozzle with the piezoelectric injector, some Iso-SOI tests were carried out The chosen SOI was -15.5 CAD ATDC, which was the one at which the 20/80 mass distribution had an IMEP of 4.0 bar Figure represents the complete pollutant emission results obtained from the Iso-SOI tests This graph presents all tested dwell times and mass distributions With increasing dwell time, combustion noise was significantly reduced because the combustion started more smoothly due to retardation of the center of the combustion NOX formation was greatly reduced by retarding the SOI However, there was a significant increase in HC, smoke and CO emissions, mainly in the mass distributions with smaller pilot injections When the pilot fuel mass was increased (see Figure 8), the center of the combustion (center of combustion is the crank angle of 50% of heat release) was advanced towards the TDC However, this did not result in a higher IMEP The in-cylinder pressure rose earlier for longer pilot injections but also decreased earlier and remained lower during expansion The IMEP was kept constant for all mass distributions In Figure 8, the region where the pressure lines cross is shown by the red circle This effect was attributed to a completely premixed combustion, and changing the mass distributions did not deteriorate the combustion process For longer pilot injections, the RoHR is steeper, leading to an increase in combustion noise Larger pilot injections advanced the entire combustion However, it is important to point out that the distance between the peaks et al of the RoHR for the mass distributions of 20/80 and 50/50 was less than CAD The peak in-cylinder pressure also increased with larger pilot injections This led to higher NOX emissions However, the 50/50 mass distribution exhibited a reduced slope in the HRL curve CONCLUSIONS In this research, two different injection systems (based on a solenoid and a piezoelectric injector) were investigated Preliminary tests were performed to select the best conditions for the main study In both cases, the chosen injection pressure was 900 bar and the EGR rate was 45% These values were chosen to work with a pilot injection smaller than 50% of the injected mass per cycle Different injection strategies were tested for each injection system There were two injection events per cycle The ISFC was kept constant at 250 g/kW.h during the tests The injection strategies were characterized by sweeping the mass distributions and dwell times A small sequence of tests was executed at constant SOI, in order to study the isolated effects of the mass distribution for different dwell times or combustion events Based on the results from the preliminary test and engine test, it is possible to conclude: (1) Independently of the injection system, 900 bar was the most suitable injection pressure for this engine The lowest injection pressure of 600 bar resulted in higher HC and smoke emissions This could have been due to a longer fuel atomization process than the other injection pressures Higher injection pressures presented lower smoke but higher combustion noise This could be attributed to the fact that more fuel mass is injected before the combustion process is started, and the atomization of the fuel is better at higher injection pressures A 40% EGR rate presented excessive combustion noise and high NOX emissions due to the fast premix combustion and high in-cylinder temperature On the other hand, a 50% EGR rate caused excessive reduction of O2 concentration, reduced NOX and combustion noise, 146 J BENAJES but the HC and CO emissions increased to unacceptable levels (2) With the injection pressure at 900 bar, mass distributions with the first injection larger than 50% of the injected mass present high CO and HC emissions However, it has been observed that the use of a postinjection smaller than 40% of the total injected mass significantly reduces smoke formation (3) An increase in dwell time with constant SOI produced a smoother start of combustion and a cooler overall combustion process, reducing noise and NOX emissions However, the emissions of HC and CO increased under these conditions (4) Increasing the pilot injection quantity caused an increase in combustion noise and NOX emissions due to faster, hotter premixed combustion The opposite effect was observed when the dwell time between injection events was increased (5) The solenoid injection system presented unsatisfactory results due to high HC and CO emissions independent of the number of nozzle orifices However, the 10orifice nozzle resulted in levels of combustion noise close to the target, while NOX remained under the limit Smoke emission increased and stayed close to the limit (6) The piezoelectric injection system with 6-orifice and 12-orifice nozzles, presented unacceptable results in terms of HC and CO emissions The combustion noise targets were achieved using both nozzles with dwell times around 1.0 ms Moreover, the 6-orifice nozzle showed higher smoke levels than the 12-orifice nozzle (7) For the same engine test configuration, the solenoid injector presented a slightly retarded combustion process compared to the piezoelectric injector due to the longer time needed for it to open completely This behavior slightly changed the in-cylinder temperatures, favoring NOX formation before the maximum BGT was reached and soot oxidation at the end of the combustion process (8) The piezoelectric injection system presented better results in terms of pollutant emissions It also permitted more accurate control of the injection parameters, including the possibility of injecting very small quantities of fuel in each injection event REFERENCES Akihama, K., Takatori, Y and Inagaki, K (2001) Mechanism of the smokeless rich diesel combustion by reducing temperature SAE Paper No 2001-01-0655 Benajes, J., Molina, S., De Rudder, K M and Ben Hadj Hamouda, H (2006) The use of micro-orifice nozzles and swirl in a small HSDI engine operating at a late split-injection LTC engine J Automobile Engineering, et al 220, 1807−1816 Benajes, J., Molina, S., De Rudder, K and Amorim, R (2007) Optimization toward low temperature combustion in a HSDI diesel engine, using consecutive screenings SAE Paper No 2007-01-0911 Desantes, J., Arrègle, J., López, J and Garcia, A (2007) A comprehensive study of diesel combustion and emission with post-injection SAE Paper No 2007-01-0915 Desantes, J., Benajes, J., Molina, S and Gonzales, C (2004) The modification of fuel injection rate in heavy-duty engines Part 2: Effects of combustion Applied Thermal Engineering, 24, 2715−2726 EurActive (2004) Euro Emissions Standard for Cars Retrieved 2007, from EurActive.com: http://www.euractiv com/en/transport/euro-5-emissions-standards-cars/ article-133325 Han, Z., Uludogan, A and Hampson, G R (1996) Mechanism of soot and nox emission reduction using multipleinjection in a diesel engine SAE Paper No 960633 Hotta, Y., Inayoshi, M and Nakakita, K (2005) Achieving lower exhaust emissions and better performance in an HSDI diesel engine with multiple injection SAE Paper No 2005-01-0928 Lapuerta, M., Armas, O and Hernandez, J (1999) Diagnosis of DI diesel combustion from in-cylinder pressure signal by estimation of mean thermodynamic properties of gas Applied Thermal Engineering, 19, 513−529 Montgomery, D., Chan, M., Chang, C., Farrell, P and Reitz, R (1996) Effect of injector nozzle hole size and number on spray characteristics and the performance of heavy duty D.I diesel engine SAE Paper No 962002 Nehmer, D and Reitz, R (1994) Measurement of the effect of injection rate and split injections on diesel engine soot and NOx emissions SAE Paper No 940668 Pierpont, D., Montgomery, D and Reitz, R (1995) Reducing particulate and NOx using multiple injections and EGR in a D.I diesel SAE Paper No 962002 Robert Bosch GmbH (2004) Diesel-Engine Management 3rd edn SAE Warrendale PA The European Commission (2006a) Euro and will Reduce Emissions from Cars Retrieved 2007, from EUROPA: http://europa.eu/rapid/pressReleasesAction do?reference=MEMO/06/409&format=HTML&aged= 0&language=EN&guiLanguage=en The European Commission (2006b) Tighter Wmission Limits for Cars After EP Adoption of Euro and Retrieved 2007, from EUROPA: http://europa.eu/rapid/pressReleases Action.do?reference=IP/06/1800&format=HTML&aged =0&language=EN&guiLanguage=en Tow, T., Pierpont, D and Reitz, R (1994) Reducing particulate and NOx emissions by using multiple injections in a heavy duty D.I diesel engine SAE Paper No 940897 International Journal of Automotive Technology, Vol 11, No 2, pp 147−153 (2010) DOI 10.1007/s12239−010−0020−6 Copyright © 2010 KSAE 1229−9138/2010/051−02 DESIGN OF ACTIVE SUSPENSION AND ELECTRONIC STABILITY PROGRAM FOR ROLLOVER PREVENTION S YIM , Y PARK and K YI 1)* 1) 2) 3) BK21 School for Creative Engineering Design of Next Generation Mechanical and Aerospace Systems, Seoul National University, Seoul 151-742, Korea Department of Mechanical Engineering, KAIST, Daejeon 305-701, Korea School of Mechanical and Aerospace Engineering, Seoul National University, Seoul 151-742, Korea 2) 3) (Received November 2008; Revised September 2009) ABSTRACT−This paper presents a method for the design of a controller for rollover prevention using active suspension and an electronic stability program (ESP) Active suspension is designed with linear quadratic static output feedback control methodology to attenuate the effect of lateral acceleration on the roll angle and suspension stroke via control of the suspension stroke and tire deflection of the vehicle However, this approach has a drawback in the loss of maneuverability because the active suspension for rollover prevention produces in vehicles an extreme over-steer characteristic To overcome this drawback of the active suspension based method, ESP is designed Through simulations, the proposed method is shown to be effective in preventing rollover KEY WORDS : Rollover prevention, Active suspension, ESP, Optimal static output feedback control, Maneuverability NOMENCLATURE ay g m ms mu Ix Iz ks kt bs h hs Cf Cr tf lf lr r KB vx Kg ed the number of rollover accidents For example, in the USA, there have been 4,045 fatalities and 88,000 injuries caused by non-collision rollover accidents in 2004 (NHTSA, 2004a) Most rollover accidents are fatal For instance, in 2003, the portion of rollovers in all crashes was approximately 3%, but 33% of all fatalities were caused by rollovers (NHTSA, 2003) As shown in Figure 1, the factors influencing rollovers are the lateral acceleration ay, the distance from the roll center to the center of gravity hs, and the lateral tire force Fy The untripped rollover occurs due to a large lateral acceleration generated by excessive steering at high speed On a low-friction road or at low speed, the rollover cannot occur because of insufficient lateral acceleration or lateral tire force Based on this observation, to prevent rollovers, it is necessary to reduce the effect of the lateral acceleration and the lateral force on vehicles Following the aforementioned idea, several control schemes were proposed to prevent rollovers The most common scheme is to reduce the lateral acceleration through decreasing a reference yaw rate with differential braking or active front steering to produce a vehicle with under-steer characteristics (Odenthal et al., 1999; Chen and Peng, 2001; Ungoren and Peng, 2004; Yoon et al., 2006; Schofield and Hagglund, 2008) However, this approach has the drawback of deteriorated maneuverability, and the yaw rate tracking performance due to this loss of maneuverability may cause another accident such as a crash or tripped rollover Another approach for rollover prevention is to control the lateral load transfer with an active suspension, which : lateral acceleration : gravitational acceleration constant : vehicle total mass : sprung mass : unsprung mass : roll moment of inertia about roll axis : yaw moment of inertia about yaw axis : stiffness of a suspension spring : stiffness of a tire : damping coefficient of a suspension damper : height of C.G from ground : height of C.G from a roll center : cornering stiffness of a front tire : cornering stiffness of a rear tire : front track width : distance from C.G to a front axle : distance from C.G to a rear axle : radius of a wheel : pressure-force constant : longitudinal velocity of a vehicle : gain of sliding mode controller INTRODUCTION Over the last decade, a widespread supply of SUVs (Sports Utility Vehicles) with high centers of gravity (C.G.) increas*Corresponding author e-mail: thewait@naver.com 147 272 M S KIM, Y G MOON, G D KIM and M C LEE Figure Schematic diagram for sensations of motions in actual vehicle Figure Schematic diagram for sensations of motions by the classic washout algorithm Figure Schematic diagram for sensations of motions by the proposed PRSM based washout algorithm gravity in a driver’s seat coordinate system The specific force in a driver’s seat coordinate system is represented as (6) Figure shows a schematic diagram for sensation of motions by a classic washout algorithm Figure shows the schematic diagram for sensation of motions by the proposed washout algorithm 4.1.2 Results of the simulation For evaluating the proposed washout algorithm, we acquired vehicle dynamics data through various driving scenarios The scenarios are composed of rapid acceleration-deceleration, uphill-downhill way, slalom, and right-left turn The cut off frequency of the high-pass filter and low pass filter was 0.01 Hz We compared the sensation without washout, with classic washout and with the proposed washout Results of the simulation are shown in Figures Figure Motion sensation in the case of rapid acceleration (deceleration) through 11 Figure shows translational motion sensation in the case of rapid acceleration and deceleration In parts of acceleration and deceleration, we can see that the proposed algorithm has better performance than others Figure 10 shows translational motion sensation in the case of an uphill and downhill road As stated above, the proposed washout algorithm has superior performance on uphill and downhill roads because the PRSM simulates the gradient of the road Figure 11 shows rotational motion sensation in the case of slalom driving Because it is difficult to show the performance difference in a graph, we calculated the correlation coefficient between the sensation in the case of without washout and the sensation in the case of classic washout In addition, we calculated the correlation coefficient between the sensation without washout and the sensation of the proposed washout If the correlation coefficient is near to the constant 1, it means that the performance of the washout algorithm was good Table shows the correlation coefficients In most cases, the proposed washout algorithm has better performance than the classic washout algorithm, in only the tilt coordination washout algorithm, and in the case of only PRSM We know that the rotational motion sensation in the case of only the tilt coordination washout algorithm is worse PARTIAL RANGE SCALING METHOD BASED WASHOUT ALGORITHM FOR A VEHICLE Figure 10 Motion sensation in the case of an uphilldownhill road 273 Table Correlation coefficient with without washout New tilt Classic Only new tilt Only coordination washout coordination PRSM algorithm algorithm algorithm and PRSM Left-right turn Surge 0.7683 0.8802 0.7740 0.8863 Sway 0.7550 0.8922 0.8131 0.9078 Roll 0.7163 0.3433 0.9939 0.8126 Pitch 0.4544 0.1412 0.9815 0.4775 Rapid acceleration-deceleration Surge 0.7682 0.9364 0.6967 0.9403 Sway 0.5485 0.8582 0.6132 0.8782 Roll 0.9251 0.8299 0.9660 0.9392 Pitch 0.5261 0.1634 0.9839 0.4571 Uphill- downhill road Surge 0.7000 0.8567 0.8163 0.9352 Sway 0.8163 0.9288 0.8579 0.9375 Roll 0.8255 0.6737 0.9793 0.9028 pitch 0.5520 0.5061 0.9716 0.8621 Slalom Surge 0.7986 0.9290 0.7725 0.9333 Sway 0.9829 0.9887 0.9897 0.9909 Roll 0.7729 0.6821 0.9967 0.9661 pitch 0.5359 0.1867 0.9825 0.5342 than one of the classic washout algorithm 4.2 Simulator Sickness Survey A survey was also conducted for the qualitative evaluation of the proposed washout algorithm The questionnaire consisted of questions about motion sensation and ones about simulator sickness Subjects could choose answers among very different (1), different (2), normal (3), similar (4), equal (5), and very equal (6) in motion sensation questions and from none (1) to very bad (6) in simulator sickness questions In the case of the proposed algorithm, subjects answered Figure 11 Motion sensation in the case of slalom driving Figure 12 Average values of motion sensation 274 M S KIM, Y G MOON, G D KIM and M C LEE stress Thus, the variation pattern of RRV is also useful to check the stress index In general, activation of the sympathetic nervous system in response to stress produces the low frequency (LF) variation pattern of RRV, while activation of the parasympathetic nervous system in response to comfort produces the high frequency (HF) variation pattern of RRV We get the stress index of subjects by calculating the ratio of LF parts to HF parts of RRV The higher the ratio of LF to HF is, the greater the stress is In this experiment, the results obtained show that the stress index is decreased in 13 subjects among 16 Table shows the result of the ECG analysis Figure 13 Average values of simulator sickness that the motion sensation was improved to 21% and simulator sickness was improved to 35% Furthermore, the majority of subjects answered that the time delay phenomenon decreased Figure 12 shows the average values of motion sensation, and the average values of the simulator sickness are shown in Figure 13 4.3 Bio-signal Analysis In this paper, we adopt the electrocardiogram (ECG) analysis method to evaluate the proposed washout algorithm because the survey can produce inaccurate results due to subjects’ wrong memories, faithfulness, and etc The RR interval (RRV) of the electrocardiogram is useful to evaluate performance of an autonomic nerve, which is part of both the sympathetic and para-sympathetic nervous system The variation pattern of RRV depends on an activation of an autonomic nerve that is sensitive to Table Ratio of stress decrement LF/HF No of Classic Proposed subjects washout washout algorithm algorithm 4.255 3.573 1.855 1.745 2.710 2.457 2.651 2.049 3.193 2.597 5.068 3.513 4.866 2.224 2.412 2.227 14.673 13.115 10 4.325 3.827 11 0.886 0.725 12 1.010 0.612 13 3.530 2.186 Average 3.957 3.142 Ratio of stress decrement 16.0 5.9 9.3 22.7 18.7 30.7 54.3 5.6 10.6 11.5 18.2 39.4 38.1 21.6 CONCLUSION In this paper, we proposed the partial range scaling method (PRSM) based washout algorithm, which consists of the PRSM and the new tilt coordination algorithm To overcome the phase delay and the signal distortion of the classic washout algorithm, we eliminated the low pass filter in tilt coordination algorithm and proposed PRSM using the roll and pitch angle of vehicle dynamics The PRSM is an efficient method to give a driver improved rotational motion sensation and restrict the motion of the VDS within the kinematic limit The proposed algorithm was evaluated by computer simulation, simulator sickness survey, and ECG analysis In the computer simulation using the human perception model, it was verified that the proposed washout algorithm decreased the time delay and signal distortion caused by filters In the survey, the majority of the subjects answered that they felt better sensation and less simulator sickness than the classic washout algorithm Finally in the ECG analysis, the stress indexes of 81% of subjects decreased to 78.4% compared with that of the classic washout algorithm It was verified by three methods that the proposed washout algorithm for VDS is more efficient than classic algorithm REFERENCES Drosdo, J and Panik, F (1985) The Daimler Benz driving simulator, A tool for vehicle development SAE Paper No 8503345 Freeman, J S., Wastson, G., Papelis, T E., Lin, T C., Tayyab, A., Romano, R A and Kuhl, J G (1995) The iowa driving simulator: An implementation and application overview SAE Paper No 950174 Grant, P R and Reid, L D (1997) Motion washout filter tunning: Rules and requirements AIAA Flight Simulation Technologies 34, 2, 145−151 Kim, M S., Lee, S Y and Yu, S B (2007) Development of a vehicle simulator based testing method for telematics software development SAE Paper No 2007-01-0945 Parrish, R V., Dieudonne, J E., Bowles, R L and Martin, PARTIAL RANGE SCALING METHOD BASED WASHOUT ALGORITHM FOR A VEHICLE D J (1975) Coordinated adaptive washout for motion simulators J Aircraft 12, 1, 44−50 Peter, C C and Burnell, T M (1981) Analysis Procedures and Subjective Flight Results of a Simulator Validation and Cue Fidelity Experiment NASA Technical Memo- randum 88270 Sivan, R., Ish-shalom, J and Huang, J K (1982) An optimal control approach to the design of moving flight simulators IEEE Trans Systems, Man and Cybernetics SMC-12, 6, 818−827 Suetomi, T., Horiguchi, A., Okamoto, Y and Hata, S 275 (1991) The driving simulator with large amplitude motion system SAE Paper No 910113 You, K S., Lee, M C., Kang, E G and Yoo, W S (2005) Development of a washout algorithm for a vehicle driving simulator using new tilt coordination and return mode J Mechanical Science and Technology 19, 1, 272−282 Yu, S B., Lee, S Y and Kim, M S (2007) Development of a virtual reality based vehicle simulator system for test and development of ASV, telematics and ITS SAE Paper No 2007-01-0946 Copyright © 2010 KSAE 1229−9138/2010/051−17 International Journal of Automotive Technology, Vol 11, No 2, pp 277−282 (2010) DOI 10.1007/s12239−010−0035−z SIMPLE DESIGN APPROACH FOR IMPROVING CHARACTERISTICS OF INTERIOR PERMANENT MAGNET SYNCHRONOUS MOTORS FOR ELECTRIC AIR-CONDITIONER SYSTEMS IN HEV * S I KIM, G H LEE, J J LEE and J P HONG Department of Automotive Engineering, Hanyang University, Seoul 133-791, Korea (Received 18 February 2009; Revised 20 July 2009) ABSTRACT−In this paper, a simple design method for improving the performance of an interior permanent magnet synchronous motor (IPMSM), for driving the air-conditioning compressor used in hybrid electric vehicles, is presented There are many design methods that optimize the IPMSM Each method deals with a variety of design factors, such as slot opening, pole arc, and rotor shape However, as the number of design variables increases, a lot of modeling and analysis time is needed in order to improve the characteristics of an IPMSM This paper demonstrates that the optimization of a double-layer IPMSM, satisfying the given design conditions, is possible with only a flux barrier shape design Then, response surface methodology is applied as the optimization method, and the validity of the design approach is verified by comparison with test results KEY WORDS : Air-conditioning compressor, IPMSM, Optimization, Response surface methodology to many design factors and the interactions between them Unfortunately, there are very few technical papers that discuss how variations of design variables effect torque ripple, cogging torque and harmonics of back-EMF The small number of papers that have appeared (Kioumarusi , 2006; Sanada , 2004) offer a design method for mitigating overall torque pulsation under various load conditions The goal of this paper is to present a relatively simple and feasible design approach that will facilitate an improvements in the above mentioned characteristics without any sacrifice to other performance characteristics of the doublelayer IPMSM The method employs the response surface method (RSM) and considers the flux barrier as the only design factor RSM is well suited for making empirical models that relate the performance of a motor to the design parameters With these empirical models, the objective functions with restraints are easily created and a lot of computational time can be saved (Jeon , 2006; Qinghua , 2004) Furthermore, an obtained response surface INTRODUCTION Idling stop systems to reduce fuel consumption are of interest for developers of hybrid electric vehicles (HEV) In order to adapt to the idling stop system, the automobile airconditioner must continue to work even while the engine is stopped As such, conventional engine-driven compressors are gradually being replaced with electric motor-driven models The motors for driving air-conditioning compressors in HEVs must be miniature, lightweight, and highly efficient Interior permanent magnet synchronous motors (IPMSMs) are well suited to these design requirements due to their high power density, high efficiency and wide speed region (Murakami , 2001) However, IPMSMs have limitations related to torque characteristics Torque ripple and cogging torque are relatively large compared with a surface permanent magnet synchronous motor These limitations are mainly the result of discontinuous reluctance variation between the rotor and stator (Kioumarusi , 2006; Bianchi , 2006; Islam , 2005; Sanada , 2004) Especially, because of d-axis current to utilize reluctance torque and increase the motor speed, the back-electromotive force (back-EMF) of the IPMSM operated in wide speed range contains many harmonics, and the torque ripple rises greatly when driven by a sinusoidal current (Lee , 2008) Therefore, it is necessary to optimize the shape of the IPMSM in order to improve torque performance This optimization is very complex and difficult due et al et al et al et al et al et al et al et al et al et al * Corresponding author e-mail: hongjp@hanyang.ac.kr Figure Configuration of initial designed model 277 278 S I KIM, G H LEE, J J LEE and J P HONG Table Dimension and specifications of double-layer IPMSM for electric air-conditioner system Items Value Stator outer diameter 117.2 mm Rotor outer diameter 70.8 mm Stack length 15 mm Air-gap 0.6 mm Br (@20~25oC) 1.22~1.28 T Number of poles Cooling method Refrigerant cooling DC link voltage 155 V Maximum terminal voltage 98.6 V Rated output power kW Maximum current 17 Arms Base, Maximum speed 3500, 7500 rpm supplies a designer with a description of system responses based on the behavior of design parameters within a design space Conventional optimization methods cannot provide this In the end, the utility of the method for the flux barrier shape design is verified by test results CHARACTERISTICS OF INITIAL MODEL vq = Ra + pLd ω Ld –ω L q R a + pLq id iq + id ωψa (1) iq = [ ψa iq + ( Ld – Lq ) id iq ] T Pn Lq according to the current and to β and q as a function of current and β are shown in Figure In (3), ψa and ψo are fundamental components calculated by Fourier analysis The steady-state phasor diagram for the IPMSM is shown in Figure (Morimoto , 1990) ψ0cosα – ψa L =ψ sin α -(3) L = -et al d id q iq where, ψo is the total flux linkage considering the armature reaction effects, and α is the phase difference between ψa and ψo The characteristics of the initial model are predicted with d and q estimated as in (1) and (2) The following limitations on armature current and terminal voltage were considered: L L = Va = id + i2q ≤ Iam vd + v2q ≤ Vam (4) (5) (2) where, d and q are components of armature current, d, q are components of terminal voltage ψa 3/2 ψf ; ψf : is the maximum flux linkage of the permanent magnet, a is the armature winding resistance, d, q are inductance along d-, and q-axis, = / , and n is the number of pole pairs At the base and maximum speed, input armature current and current angle (β ) are required to estimate torque ripple by FEA (Kioumarusi , 2006; Sanada , 2004) To get these values, d and q should be computed according to the change of armature current and β In this paper, they are obtained by FEA cubic spline interpolation and (3) d i and Figure Phasor diagram of the IPMSM Ia =PnψaIacosβ + - ( Ld – Lq) Ia2sin2β Ld L Figure shows the initial configuration of the double-layer IPMSM designed for driving the compressor of a HEV The constant power speed range (CPSR) of the initial model is 3500 rpm to 7500 rpm The main dimensions and specifications are listed in Table The characteristics of the model are estimated by finite element analysis (FEA) and the voltage and torque equation; mechanical and iron loss are ignored The equations in normal operation are expressed in d-q coordinates as follows (Chin and Soulard, 2003): vd Figure i v v R L p d dt P et al L L et al L L Figure Speed versus torque and output performance of the initial model SIMPLE DESIGN APPROACH FOR IMPROVING CHARACTERISTICS OF INTERIOR PERMANENT MAGNET 279 where, am and am are peak values of current and voltage, respectively The entire torque-speed operation region, considering the above control conditions, is acquired in the following manner In the anterior region of base speed, maximum torque per ampere control is used In the posterior region flux, weakening control is applied Characteristics of the initial model, such as speed versus torque (Figure 4), torque waveforms at the base (Figure 5), and maximum speed and cogging torque estimated by FEA (Figure 6), are shown Input currents were 15.3 A and 12.5 A, and β was 32.5o and 63.6o, respectively The back-EMF of the initial model and its total harmonic distortion (THD) are given in Figure I V Figure Back-EMF and THD of initial model at 3500 rpm DESIGN OPTIMIZATION 3.1 Selection of Design Factors Section demonstrated that the initial design of the IPMSM satisfied the given design conditions In the IPMSM, the operating constraints on input current, terminal voltage, and CPSR critically depend on motor parameters, such as the flux linkage by the permanent magnet, as well as d- and q-axis inductance (Morimoto and Takeda, 2000) Therefore, in the initial model, the size and position of permanent magnet and air-gap length cannot be changed Due to the fill factor and auto winding, the slot opening, the teeth, and the yoke width also should not be altered As a result, the only design factor selected in this paper is the flux barrier angle Figure 8, a magnified view of the region of Figure Figure Torque waveforms of initial model at the base and maximum speed Figure Design variables of the initial model surrounded by a dotted line, indicates the factors In the shape optimization there are many limitations that make it difficult to consider all the design variables 3.2 Application of RSM RSM is a set of statistical and mathematical techniques to find the “best fit” response of a physical system through experiment or simulation RSM has recently been recognized as an effective approach for modeling the performance of electrical devices In RSM, a polynomial model, called a fitted model, is usually constructed to represent the relationship between performance and design parameters In this paper, RSM is used to make appropriate response models of torque ripple, cogging torque and THD of back-EMF of the initial model A quadratic approximation function of the model is commonly used to construct the fitted response surface In general, the response model can be written as follows: Y= b0 + k k i=1 i i+ i=1 ii i + i≠j ij i j + ε (6) where, o…ij are regression coefficients for design variables, and ε is a random error In this paper, the least squares method is used to estimate unknown coefficients The fitted coefficients and response model can be written as: (7) b X Y –1 X T Y (8) Y =Xb Where, is matrix notation for the levels of the indepenb ′= ( ′ Figure Cogging torque waveform of initial model k ∑bx ∑b x ∑b x x ′ ) ′ X 280 S I KIM, G H LEE, J J LEE and J P HONG Table Design space of CCD Levels of design factor Design factors −α −1 o 60 90 120 θ1 [ ] 39.5 θ2 [ o ] 43.2 50 60 70 θ3 [ o ] 36.6 40 45 50 Table FEA results based on CCD θ1 [ o ] θ2 [ o ] θ3 [ o ] YAT YTr_B YTr_M 60 50 40 5.65 9.56 28.90 120 50 40 5.69 10.54 25.0 60 70 40 5.70 13.51 32.85 120 70 40 5.74 14.60 27.03 60 50 50 5.68 12.50 32.30 120 50 50 5.73 12.57 29.78 60 70 50 5.76 15.23 36.03 120 70 50 5.80 15.86 31.34 39.5 60 45 5.69 11.07 31.45 140.4 60 45 5.77 10.75 24.49 90 43.2 45 5.67 10.41 28.57 90 76.8 45 5.62 14.41 31.52 90 60 36.6 5.62 14.23 31.47 90 60 53.4 5.75 15.83 35.07 90 60 45 5.74 12.89 30.25 a 140.4 76.8 53.4 YCogT 0.084 0.087 0.150 0.055 0.090 0.088 0.156 0.055 0.140 0.176 0.054 0.087 0.087 0.092 0.091 YTHD 5.39 4.84 4.57 3.61 3.96 3.32 4.04 3.05 4.41 4.08 4.38 5.01 5.0 3.55 4.22 Table Fitted coefficients and response models Coeffi′Tr_B ′Tr_M ′CogT ′AT cients β′0 3.8020 35.75 85.39 −0.8250 β′1 −0.0002 0.1690 0.1009 0.0007 β′2 0.0270 0.4860 0.4458 0.020 β′3 0.0410 −2.3710 −3.584 0.0118 β′11 −0.0007 −0.0009 β′22 −0.0003 −0.0011 −0.0008 −0.0001 β′33 −0.0005 0.0325 0.0424 −0.0001 β′12 0.0003 −0.0017 −0.0001 β′13 −0.0011 0.0021 β′23 0.0001 −0.0050 −0.0017 Y Y Y Table Optimal condition Design factors Initial model o θ1 [ ] 60 θ [ o] 60 θ [ o] 45 Y ′THD Y 19.50 0.0262 −0.3102 −0.1928 −0.0001 0.0010 −0.0019 −0.0003 −0.0001 0.0047 Optimized model 140.4 52.2 45 *YAT: average torque at the base speed; YTr_B and YTr_M : torque ripple at the base and maximum speed; YCogT: peak to peak of cogging torque; YTHD : THD of back-EMF dent variables, is the transpose of the matrix , and is the vector of the observations Central composite design (CCD) was used as the experimental design method, to estimate the fitted model of each response (Montgomery, 2001) CCD consists of three portions: 1) a complete factorial design in which the factor levels are coded into –1 and 1, 2) axial points at a distance α from the center point, and 3) one design center point Table shows the design area of CCD The FEA results, based on CCD, are listed in Table Input current and current angle at the base and maximum speed are the same those of the initial IPMSM Finally, the fitted models, obtained by the results of Table 3, (7) and (8) are shown in Table The flux barrier design with the fitted models was performed The design objectives and constraints are as follows: • Design objectives: YTr_B 10% , YTr_M 30% , YCogT 0.16 Nm , YTHD 4% • Subject to: YAT 5.5 Nm , Output power ≥ kW Table shows the optimal point satisfying the design X T X k ′ ≤ ′ ′ ≤ ′ ′ Y Figure Responses of each fitted model according to the design factors objectives, and Figure describes the results of each fitted model corresponding to the point Each model response varies greatly according to the variation of design factors This means that it is difficult to find the optimal condition that minimizes every response Thus, a proper trade-off is required, according to the application field of the IPMSM ≤ ≤ ≥ Figure 10 Optimized model SIMPLE DESIGN APPROACH FOR IMPROVING CHARACTERISTICS OF INTERIOR PERMANENT MAGNET 281 TEST RESULTS The optimized double-layer IPMSM was fabricated as shown in Figure 10 and tested in order to verify the validity of the results acquired through the optimization method Figure 11(a) shows a testing apparatus for measuring torque ripple Torque ripple results are displayed in Figures 11(b) and (c) The voltage amplitude obtained was Nm per V The torque ripple of the optimized IPMSM, at the load condition of base and maximum speed, was measured as 7.82% and 12.55%, while the FEA results were 10.0% and 24.7%, respectively The difference between the test results and the FEA results may have been caused by the influence of the motor and reduction gear inertia The measured and estimated cogging torques are given in Figure 12 The measured magnitude and wave-shape of the cogging torques are quite similar to the FEA predicted results Back-EMF, measured and calculated at 3500 rpm, and its THD are shown in Figure 13 The waveforms of the measured and computed back-EMF are slightly different We hypothesize that the difference is created by the manufacturing error generated in the barrier angle, slot opening, and tooth tip In the end, the accuracy of the predicted values, obtained by each fitted response model, were proved through the test and FEA results (Figure 9) CONCLUSION In this paper, RSM was presented as an optimization method for the flux barrier shape design to improve the Figure 11 Test and FEA results with respect to the torque ripple of the optimized model Figure 12 Test and FEA results with respect to the cogging torque of the optimized model 282 S I KIM, G H LEE, J J LEE and J P HONG IEEE Trans Veh Technol , , 1102−1111 Chin, Y K and Soulard, J (2003) Modeling of iron losses in permanent magnet synchronous motors with fieldweakening capability for electric vehicles Int J Automotive Technology , , 87−94 Islam, M S., Mir, S., Sebastian, T and Underwood, S (2005) Design considerations of sinusoidally excited permanentmagnet machines for low-torque-ripple applications IEEE Trans Ind Applicat , , 955−962 Jeon, M H., Kim, D H and Kim, C E (2006) Optimum design of BLDC motor for cogging torque minimization using genetic algorithm and response surface method J Electrical Engineering & Technology , , 466−471 Kioumarsi, A., Moallem, M and Fahimi, B (2006) Mitigation of torque ripple in interior permanent magnet motors by optimal shape design IEEE Trans Magn , , 3706−3711 Lee, G H., Kim, S I., Hong, J P and Bhan, J H (2008) Torque ripple reduction of interior permanent magnet synchronous motor using harmonic injected current IEEE Trans Magn , , 1582−1585 Montgomery, D C (2001) Design and Analysis of Experiments John Wiley & Sons New York Morimoto, S., Takeda, Y and Hirasa, T (1990) Current phase control methods for permanent magnet synchronous motors IEEE Trans Power Electron , , 133−138 Morimoto, S and Takeda, Y (2000) Machine parameters and performance of interior permanent magnet synchronous motors with different permanent magnet volume Elec Eng Japan , , 1403−1408 Murakami, H., Kataoka, H., Honda, Y., Morimoto, S and Takeda, Y (2001) Highly efficient brushless motor design for an air-conditioner of the next generation 42V vehicle Conf Rec IEEE-IAS Annu Meeting, 461−466 Qinghua, L., Jabbar, M A and Khambadkone, M (2004) Response surface methodology based design optimization of interior permanent magnet synchronous motors for wide-speed operation Proc PEMD, 546−551 Sanada, M., Hiramoto, K., Morimoto, S and Takeda, Y (2004) Torque ripple improvement for synchronous reluctance motor using an asymmetric flux barrier arrangement IEEE Trans Ind Applicat , , 1076−1082 55 4 41 4 42 11 Figure 13 Back-EMF measured and computed at 3500 rpm and THD characteristics of initially designed double-layer IPMSM Performance variations according to changes of design factors were easily predicted by the method, and the optimal condition to satisfy multiple design objectives was detected However, the optimal points of each response could not minimize torque ripple at the base, torque ripple at maximum speed, cogging torque, and THD of back-EMF simultaneously Moreover, the optimal conditions for torque ripple conflict with the optimal conditions for cogging torque Thus, RSM is a useful method for finding an appropriate trade-off Finally, the validity of the design method was verified by comparison with test results of the fabricated optimal model REFERENCES Bianchi, N., Pre, M D and Bolognani, S (2006) Design of a fault-tolerant IPM motor for electric power steering 44 31 40 Copyright © 2010 KSAE 1229−9138/2010/051−18 International Journal of Automotive Technology, Vol 11, No 2, pp 283−287 (2010) DOI 10.1007/s12239−010−0036−y ROLLOVER MITIGATION FOR A HEAVY COMMERCIAL VEHICLE Y I RYU , D O KANG , S J HEO and J H IN 1) 1) 2)* 3) Graduate School of Automotive Engineering, Kookmin University, Seoul 136-702, Korea School of Mechanical and Automotive Engineering, Kookmin University, Seoul 136-702, Korea NewTech T&M, 934-2 Koryeom-ri, Cheongbuk-myeon, Pyeongtaek-si, Gyeonggi 451-830, Korea 1) 2) 3) (Received 27 April 2009; Revised 17 August 2009) ABSTRACT−A heavy commercial vehicle has a high probability of rollover because it is usually loaded heavily and thus has a high center of gravity An anti-roll bar is efficient for rollover mitigation, but it can cause poor ride comfort when the roll stiffness is excessively high Therefore, active roll control (ARC) systems have been developed to optimally control the roll state of a vehicle while maintaining ride comfort Previously developed ARC systems have some disadvantages, such as cost, complexity, power consumption, and weight In this study, an ARC-based rear air suspension for a heavy commercial vehicle, which does not require additional power for control, was designed and manufactured The rollover index-based vehicle rollover mitigation control scheme was used for the ARC system Multi-body dynamic models of the suspension subsystem and the full vehicle were used to design the rear air suspension and the ARC system The reference rollover index was tuned through lab tests Field tests, such as steady state cornering tests and step steer tests, demonstrated that the roll response characteristics in the steady state and transient state were improved KEY WORDS : Active roll control, Air suspension, Anti-roll bar, Ride comfort, Rollover index, Rollover mitigation INTRODUCTION In this study, an ARC-based rear air suspension was designed and implemented The ARC system considered in this study locks and unlocks a solenoid valve-type actuator, which is attached between the ARB and the body frame, depending on the need for additional roll stiffness The rollover index-based vehicle rollover mitigation control scheme was used (Yoon et al., 2006) For the design of the air suspension and the ARC system, multi-body dynamic models of the suspension subsystem and the full vehicle were constructed Using these models, the location and specifications of air springs and the roll stiffness of the ARB were determined In addition, these models were used to estimate the performance characteristics of the ARC system The designed rear air suspension and ARC system were manufactured, and their performance was evaluated through lab and field tests Before air springs were used in the suspensions of heavy commercial vehicles, leaf springs were responsible for ride and roll Because the leaf spring generally has a high spring stiffness, considering the loading weight, the ride rate changes according to the loading This inhibits ride comfort especially when driving on an uneven road In addition, when a vehicle undergoes quick turning with heavy loading, the possibility of rollover is high due to insufficient roll stiffness To overcome the shortcomings of the leaf spring, a suspension system utilizing an air spring and passive anti-roll bar (ARB) has recently been implemented The passive ARB is effective for rollover mitigation due to the increase of the roll stiffness However, this increase may deteriorate ride comfort while driving on an uneven road because the ARB has a fixed roll stiffness unrelated to vehicle maneuvers One of the methods that may facilitate simultaneous maintenance of ride comfort and optimal control of the roll situation is active roll control (ARC) (Cebon et al., 1998; Daring and Martim, 1997; Stone and Cebon, 2002) When the roll stiffness needs to be increased, the ARB is forced to twist in the opposite direction of the roll This necessitates the use of a power supply for generating actuator or motor force, which increases cost, complexity, power consumption, and weight DESIGN FOR ACTIVE ROLL CONTROL SYSTEM In this study, the rear suspension of a commercial truck with a rigid axle and leaf springs was modified into an air suspension, an ARB, and linkages A rollover index was introduced for the ARC The specifications of the truck are as follows: − Gross vehicle weight: front/rear=3,570/7,115 kgf − Wheel base: 4,260 mm − Wheel tread: front/rear=1,855/1,660 mm *Corresponding author e-mail: sjheo@kookmin.ac.kr 283 284 Y I RYU, D O KANG, S J HEO and J H IN 2.1 Rollover Index Figure illustrates the mechanisms on a vehicle with a suspension system during cornering (Gillespie, 1992) A simple analytical solution for the rollover threshold is possible if the mass and roll of the axles are neglected Taking moments about the point where the right wheel contacts the ground and assuming the left wheel load has gone to zero gives: (1) Σ M =0= M a h−M g[ t/2−φ ( h – h ) ] where M = Body (or sprung) mass h = Height of the center of gravity above the ground h = Height of the roll center of gravity above the ground at the longitudinal CG location t = Tread The roll angle of the sprung mass, φ , is the roll rate, Rφ , times the lateral acceleration, α The roll rate is the rate of change of the roll angle with lateral acceleration expressed in units of radians per g Substituting to eliminate the roll angle and solving for lateral acceleration yields: a t - -1 = (2) g h [ – Rφ ( – h / h ) ] s y s r s r y y r Thus, the static roll response of a vehicle takes the form shown in Figure At low levels of lateral acceleration the vehicle roll-response increases linearly with a slope equal to the roll rate This proceeds until the inside wheel lifts off the ground When the inside wheel lifts the rollover threshold has been reached Thereafter, the roll response follows the downward sloping line Through the analysis of static rollover, we can estimate the roll angle threshold, φ , and the critical lateral acceleration, α , that may induce rollover If the roll state of the vehicle is below the wheel lift threshold the vehicle is stable, but when it is near or exceeds the wheel lift threshold the danger of rollover exists The introduced rollover index, Equation (3), can indicate the danger of rollover based on the wheel lift threshold C and C values are tuned to establish the reference rollover index for the ARC An RI of indicates wheel-lift-off th yc · · φ (t) φ + φ (t) φ ⎞ a ⎞ - +C ⎛ RI=C ⎛⎝ -· ⎠ ⎝ a -⎠ th th y φ thφ th yc φ (t) +( – C – C )⎛⎝ ⎞⎠ ( φ ( t ) ) + ( φ· ( t ) ) 2 (3) where, < C < 1, < C < 1 2.2 ARC Scheme Figure and Table show the ARC scheme used in this study Once the measured RI value, which is calculated Figure ARC scheme Figure Roll model of a suspended vehicle Table Control modes Control Definition mode ARB-ON · Actuator state-always locked ARB-OFF · Actuator state-always unlocked ARC Figure Static rollover characteristics of a suspended vehicle · Measured RI valve > Reference RI value: Actuator state - locked · Measured RI value < Reference RI value: Actuator state - unlocked Actuator ROLLOVER MITIGATION FOR A HEAVY COMMERCIAL VEHICLE 285 from sensor signals, such as lateral acceleration, roll angle, and roll rate, reaches the reference RI value for operating the ARC, the solenoid valves of the actuators are locked and the roll stiffness of the ARB is increased Three control modes, which could be selected by the driver, were established (Table 1) The reference RI value, C1 and C2, in Equation (3), were tuned through lab tests The control modes were realized during manufacturing of the electronic control unit (ECU) for the ARC (GVW) condition The location and specifications of the air springs and the roll stiffness of the ARB were determined through kinematic and compliance simulations Additionally, the component input loads, which were needed for component design, were calculated through simulations of abuse load conditions Finally, the performances of the rear air suspension and the ARC system were estimated by performing a virtual driving simulation of the full vehicle model linked with the ARC algorithm 2.3 Multi-body Dynamic Models for Virtual Simulation Figure shows the multi-body dynamics model for the rear suspension and the full vehicle in the gross vehicle weight 2.4 Design and Manufacturing of the Rear Suspension Based on the virtual simulations, the air springs were located outside vehicle to secure the adequate roll stiffness without requiring additional springs The solenoid valves of the actuators connecting the ARB to the body frame are usually set open This causes the ARB to go down due to its own weight To prevent this problem, the ARB is attached to the body frame by a hanging spring Figure shows the prototype design of the rear suspension The hanger bracket, front/rear air spring, upper/lower bracket, V-rod axle/frame bracket, and support beam, which are needed to install suspension components on the axle or frame, were designed to withstand the component input loads calculated from the simulations under the abuse load conditions LAB TEST Figure Multi-body dynamics model for virtual simulation: rear suspension (upper) and full vehicle (lower) Figure Design of the rear suspension To observe the operating condition of the rear air suspension and ARC system, a lab test was performed, as shown in Figure To replicate a roll situation, the wheel centers at both ends of the axle were fixed, a roll moment was applied by a hydraulic actuator, and a lateral acceleration signal was generated from a signal generator At the same time, the roll angle was calculated using the displacement signals between the frame and axle measured on the left and right sides of the axle The roll rate was obtained through differentiation of the roll angle The initial value of Figure Lab test setup 286 Y I RYU, D O KANG, S J HEO and J H IN Figure Lateral acceleration for the steady state cornering test Figure Field test setup reference RI was determined through tests for various roll conditions FIELD TEST Figure 10 Roll angle for the steady state cornering test For field tests in the GVW condition, a dummy weight of tons was loaded, and various sensors and an ECU were installed, as shown in Figure A steady state cornering test and step steer test were performed to evaluate the steady state and transient responses for each control mode in each roll situation Figure 10 shows that a roll angle of 1.3 degrees occurred with the ARB-ON control mode and a roll angle of 2.6 degrees occurred with the ARB-OFF control mode Thus the addition of the roll stiffness of the ARB reduced the roll angle by 50% 4.1 Steady State Cornering Test The steady state cornering tests were performed under various velocity conditions on a circle with a radius of 40 m for the ARB-OFF and the ARB-ON control modes Figures 8~10 show the results for the 40 km/h condition The steering angle and lateral acceleration for the two control modes showed similar responses (Figures and 9) 4.2 Step Steer Test Step steer tests with a velocity of 40 km/h, a steering input angle of 120 degrees, and a steering angle rising time of 0.5 seconds were performed for the ARB-ON, ARB-OFF, and ARC control modes In Figure 11, the roll angle response in the steady state shows 2.1 degrees in the ARB-OFF mode, 1.0 degrees in the ARB-ON mode, and 1.7 degrees in the ARC mode This shows that a roll angle reduction Figure Steering wheel angle for the steady state cornering test Figure 11 Roll angle for the step steer test ROLLOVER MITIGATION FOR A HEAVY COMMERCIAL VEHICLE 287 suspension and the ARC system were designed using multibody dynamic models linked with the ARC algorithm for the suspension subsystem and the full vehicle The reference roll over index was tuned through lab tests, and the performance of the manufactured rear air suspension and ARC system was evaluated through field tests, such as steady state cornering tests and step steer tests The tests demonstrated that the roll response characteristics in the steady state and transient state were improved Figure 12 Yaw rate for the step steer test effect of 19% was achieved through the ARC (Figure 11) As shown in Figure 12, the yaw rates for the ARB-OFF and ARC modes showed similar responses, and the steady state value and overshoot for the ARB-ON mode were higher than those for the other modes This means that the vehicle is relatively unstable in the ARB-ON mode compared to the other modes because more roll moment on the rear axle induces more lateral load transfer leading to reduction of the average lateral force from both tires and thus contributing to oversteer CONCLUSIONS In this study, an ARC-based rear air suspension, for rollover mitigation of a heavy commercial vehicle, was designed and manufactured A rollover index-based vehicle rollover mitigation control scheme was used The rear air ACKNOWLEDGEMENT−This work was supported by research program 2007 of Kookmin-University in Korea The authors gratefully acknowledge the financial support for research REFERENCES Cebon, D J M Sampson (1998) An investigation of roll control system design for articulated heavy vehicles Proc AVEC98, 311−316 Daring, J and Martin, T J (1997) A theoretical investigation of a prototype active roll control system Proc Instn Mech Engrs., 211, 3−12 Gillespie, T D (1992) Fundamentals of Vehicle Dynamics Warrendale PA Society of Automotive Engineers Stone, E and Cebon, D (2002) A preliminary investigation of semi-active roll control 6th Int Symp Advanced Vehicle Control, AVEC2002 Yoon, J., Yi, K and Kim, D (2006) Rollover index-based rollover mitigation control system Int J Automotive Technology 7, 7, 821−826